Engine-driven reciprocating piston compressors have been known since the industrial revolution. Compressor designs have improved over time to improve volumetric and overall energy efficiency, to improve performance for higher compression ratios and higher discharge gas pressures, to increase durability, and to reduce manufacturing costs. Improvements are still being made today.
For a compressor with a cam and roller tappet assembly, with higher compression ratios, and higher discharge pressures come the potential for higher Hertzian pressures between the tappet roller and cam. Hertzian pressure can be reduced by increasing the size of the roller to increase the contact surface area between the roller and cam. However, there are practical limits to the size of the roller because increasing the roller size also adds to the weight and overall size of the compressor. For a compactly designed high-pressure compressor, it is usually impractical to maintain Hertzian pressure below desired limits by increasing roller size alone. Higher Hertzian pressures beyond material limitations will increase wear and can result in mechanical failure and consequently reduce the service life of the rollers and/or cams if measures are not taken to reduce Hertzian pressure and/or increase the durability of the tappet roller and cam.
The goal of increasing volumetric efficiency has led to the design of compressors with low cylinder bore diameter to piston stroke ratios. Volumetric efficiency is inversely proportional to the parasitic volume, which is a physical characteristic associated with each compressor design. The parasitic volume is the gas-filled volume remaining in the compression chamber at the end of a compression stroke, when the piston is fully extended (when the piston is at top dead center). Some clearance is required between the fully extended piston and the cylinder head to avoid damage that might be caused by the piston contacting the cylinder head or contact with valve components that might be extendable into the compression chamber. The gap between the piston and the cylinder bore that is between the piston head and the first piston ring seal also contributes to the parasitic volume. There may also be respective passages between inlet and outlet valve seats and the compression chamber that also contribute to the parasitic volume. The compressor does work to compress the gas in the parasitic volume to a high pressure, but at the end of the compression stroke, the piston can not move beyond its fully extended position to discharge the compressed gas from the parasitic volume of the compression chamber. Furthermore, when the compressor piston retracts during the subsequent intake stroke to draw more gas into the compression chamber for the next compression stroke, new gas can not be drawn into the cylinder until after the compressed gas that was in the parasitic volume has expanded to the point where its pressure is less than the supply pressure of the gas that is to be drawn in through the inlet valve. Therefore, a larger parasitic volume reduces the amount of new gas that can be drawn into the compression chamber on each subsequent intake stroke and this results in lower volumetric efficiency.
For known high-pressure gas compressors it is considered necessary to reduce the cylinder bore diameter to piston stroke ratio, to reduce the parasitic volume and improve volumetric efficiency to desirable levels. That is, since there is a limit to how much one can reduce the spacing between the piston and the cylinder head at the end of the compression stroke, in modern compressors, for a given displacement, parasitic volume is reduced by reducing the size of the bore and increasing the stroke length. For example, with known high-speed piston compressors for compressing natural gas to 250 bar, bore to piston stroke ratios of the high pressure stage are normally less than 0.5 and typically as low as 0.3, which corresponds to a stroke length that is up to 3.4 times larger than the cylinder bore diameter. For low and medium compression stages, the respective bore to stroke ratios can be as high as 2 and as low as 0.5.
Mechanically driven piston compressors can use a crankshaft connected to the piston by piston rods, like the arrangement used for internal combustion engines. The compressor can even be incorporated into the engine block, using the same crankshaft that is driven by the engine pistons, with some of the pistons being used by the engine to generate power and other pistons used for gas compression, such as is disclosed in U.S. Pat. No. 5,400,751, entitled “Monoblock Internal Combustion Engine With Air Compressor Components”. However, such arrangements can be more complicated and less efficient than an arrangement that employs a piston driven by a cam and roller tappet assembly, such as that disclosed by Miller et al. U.S. Pat. No. 5,078,580. In addition, a compressor with a piston driven by a cam and roller tappet assembly can be more compact so that the size of the compressor can be reduced, compared to the size of a compressor driven by a crankshaft and a piston rod. Miller discloses a piston assembly wherein the compressor piston comprises a stem that is screwed into a crosshead. The piston assembly further comprises a roller mounted in the crosshead by a pin. A spring causes the piston to retract downwards to follow the cam surface. However, a problem with this arrangement is that the piston, crosshead, and roller are fixedly attached to each other and each of these components must be aligned with another component: the piston with the cylinder, the crosshead with a guide, and the roller with the cam. With compressors in general, and especially for compressors designed for high gas pressures, it is desirable to reduce the clearance between the piston and the cylinder. Consequently, the assembly taught by Miller would be expensive to manufacture because of the small manufacturing tolerances needed to for alignment of the piston in the cylinder, the crosshead in the guide, and the roller on the cam. Miller also does not disclose an arrangement that would be suitable for operating with longer intervals between servicing and high durability. For example, Miller does not disclose a means for lubricating the tappet roller assembly. Furthermore, another important drawback of the compressor disclosed by Miller is that it does not provide a means for reducing the force acting on the piston resulting from the gas pressure in the compression chamber and consequently the Hertzian pressure between the roller and cam can be too high. A problem specific to cam and roller tappet assemblies is wear of the cam and rollers, which is a problem that can be amplified in a compressor that is designed for handling high-pressure gases. The Hertzian pressure is the contact pressure between the cam and roller, and damage or accelerated wear can result if the Hertzian pressure is too high. Another disadvantage of excessively high forces resulting from high gas pressures in the compression chamber is that it can result in higher friction in the drive train and consequently, lower overall efficiency. For compressors with variable intake gas pressure, such as compressors that are employed to pressurize gas supplied from a storage vessel, it can be difficult to guard against excessive Hertzian pressure because gas pressure in the compression chamber is variable, depending upon gas pressure in the storage vessel.
Douville et al. U.S. Pat. No. 5,832,906 discloses an intensifier apparatus. An intensifier apparatus is a type of compressor that can be employed to increase the pressure of a gas supplied from a variable pressure source to a higher pressure. Douville discloses a two stage compressor with piping that connects the supply pipe to the back side of the first-stage piston through a back pressure port, permitting the intensifier to run in an idle operating mode with the load on the first and second stage pistons balanced while no compression takes place. Douville discloses a scotch yoke arrangement for using a rotating cam to drive the compressor pistons. Such an arrangement is useful for a two-piston, two-stage compressor but is not suitable for other arrangements, such as a single-stage, single-piston compressor, or a three-stage, three-piston compressor. Douville does not disclose a means for reducing Hertzian pressure that can be applied to each cylinder of both single and multi-piston compressors.